1. Field of the Invention
The present invention relates in general to rotary compressors and, more particularly, to a structural improvement in such rotary compressors for achieving smooth lubrication to the contact part between inner surface of a rolling piston and outer surface of an eccentric sheave of crankshaft and to the contact part between the crankshaft and bearings and thereby improving operational performance of the rotary compressors and prolonging life of the rotary compressors.
2. Description of the Prior Art
With reference to FIG. 1, there is shown a typical rotary compressor. As shown in the drawing, the typical cylindrical rotary compressor 100 includes a vertically placed crankshaft 1 which rotates in a cylinder 13 by rotational force transmitted thereto through a power transmission mechanism.
A hollow cylindrical rotor 2 is tightly fitted over the crankshaft 1 through, for example, thermal fitting, thus to be integrated with the shaft 1 into an assembly. The compressor 100 also includes a hollow cylindrical stator 3 whose inner diameter is larger that the outer diameter of the rotor 2. The stator 3 is placed about the rotor 2 such that there is a gap between the rotor 2 and the stator 3.
The inner surface of the stator 3 is provided with a plurality of longitudinal slits (not shown) of a predetermined depth, the slits being spaced apart from each other at regular intervals.
A main bearing 4 and a sub-bearing 5 are fitted over the lower section of the crankshaft 1 such that the bearings 4 and 5 are spaced apart from each other at an interval. In this case, it is typical to place the main bearing 4 above the sub-bearing 5.
The portion of the crankshaft 1 between the bearings 4 and 5 is provided with an eccentric sheave 11 as shown in FIGS. 2 and 3. In addition, a ring type rolling piston 12, having an inner diameter larger than the outer diameter of the sheave 11, is placed about the sheave 11 such that the outer surface of the sheave 11 partially contacts with the inner surface of the rolling piston 12.
The rolling piston 12 is placed in a cylinder 13. The cylinder 13 has a linear reciprocating vane 14 which is elastically biased by a spring 14a placed in the outside of the cylinder 13. The tip of the spring-biased vane 14 always contacts with the outer surface of the rolling piston 12 in the circumferential normal direction of the piston 12 so that the vane 14 linearly reciprocates during eccentric rotation of the rolling piston 12.
In the left and right sides of the vane 14, a suction port 13a and an exhaust port 13b are formed by holing the inner wall of the cylinder 13. The cylinder 13 also includes an elbow type suction pipe 6 extending to the outside of the compressor 100. The suction pipe 6 radially extends from the cylinder 13 and in turn vertically extends upward as shown in FIG. 1.
The rotary compressor 100 is totally cased by a shell 7. The inside lower section of the shell 7 is filled with oil as shown in FIG. 1. The bottom center of the crankshaft 1 is provided with an oil suction port 21 for sucking the oil of the shell 7 into the crankshaft 1. A refrigerant exhaust pipe 8 is vertically fitted into the top center of the compressor 100.
An internal oil conduit 22 longitudinally extends in the crankshaft 1 from the top center to the bottom center of the shaft 1 as shown in FIG. 3. An oil port 23 is radially formed in the middle portion of the shaft 1 such that the port 23 lets the conduit 22 communicate with the outside of the shaft 1.
The oil conduit 22 also communicates with the outside of the shaft 1 at about the top and bottom centers of the eccentric sheave 11 through radial oil ports 24 and 25. With the oil ports 24 and 25, it is possible to lubricate the contact parts between the shaft 1 and the main and sub bearings 4 and 5.
An outer hole 11a is radially formed on the outer surface of the eccentric sheave 11 as shown in FIG. 4, while a connection hole 11b extending to the outer hole 11a is radially formed in the sheave 11 as shown in FIG. 5. With the outer hole 11a and the connection hole 11b, the internal oil conduit 22 of the shaft 1 communicates with the outside of the sheave 11.
Turning to FIGS. 7, 8A and 8B, the inner surface of each of the main and sub bearings 4 and 5 fitted over the shaft 1 is provided with a lubricating groove 30. This groove 30 extends from a start point 30a to a stop point 30b. Of course, it should be understood that the two points 30a and 30b may be interchanged each other.
In the drawings, the reference numeral 18 denotes an electric power terminal.
In operation of the above rotary compressor 100, expanded gas is sucked into the cylinder 13 through the suction pipe 6 and, at the same time, the eccentric sheave 11 rotates in accordance with rotation of the crankshaft 1. Due to rotation of the sheave 1, the rolling piston 12 which is placed in the cylinder 13 and contacts with the outer surface of the sheave 11 rotates in a given direction and thereby compressing the expanded gas in the cylinder 13 so as to produce high pressure and high temperature gas. The compressed gas in turn passes through the longitudinal slits of the stator 3 and is discharged from the compressor 100 through the exhaust pipe 9 in the top of the compressor 100.
The gas compression theory of the compressor 100 will be given hereinbelow with reference to FIG. 2.
When the expanded gas has been sucked into the cylinder 13 through the suction pipe 6, the rolling piston 12 whose outer surface contacts with the tip of the spring-biased vane 14 rotates along with the eccentric sheave 11 of the shaft 1, thus to eccentrically rotate in the cylinder 13.
In this case, the rolling piston 12 rotates and revolves in the same direction as the rotating direction of the shaft 1 while the spring-biased vane 14 contacting with the outer surface of the piston 12 linearly reciprocates.
As the piston 12 eccentrically rotates in a given direction under the condition that it is applied with pushing force of the spring-biased vane 14, the gas sucked into the cylinder 13 through the suction port 13a is compressed at every rotation of the piston 12 in the cylinder 13. The compressed gas in turn is discharged from the cylinder 13 through the exhaust port 13b while overcoming the spring force of an exhaust valve spring 13c.
FIG. 6A shows pulse signals of rotation period of the rolling piston 12 and FIG. 6B is a graph showing rotation speed of the rolling piston 12 as a function of exhaust pressure, and FIG. 6C shows a device for detecting the rotation speed of the piston 12.
The rotation speed of the piston varies in accordance with the frictional force which is generated between the eccentric sheave 11 and the inner surface of the piston 12 due to radially inward biasing force resulting from the suction gas pressure, the spring force of the vane spring 14a and the exhaust gas pressure. The rotation speed also varies in accordance with the frictional force generated between the tip of the vane 14 and the outer surface of the piston 12.
As shown in FIG. 6C, a pulse is generated whenever an insulating part 13d formed on the outer surface of the rolling piston 12 meets with an electrode 19 of the rotation speed detecting device.
As shown in FIG. 6A, the rotation speed of the piston 12 becomes 132 rpm when the exhaust pressure Pd is 1.57 M and, in this case, the pulses are generated at a rate of 2.2 pulses/sec. When the exhaust pressure Pd is 2.07 Mpa, the rotation speed of the piston 12 becomes 24 rpm and, in this case, the pulses are generated at a rate of 1 pulse/about 2.5 secs.
As shown in the graph of FIG. 6B, the pressure difference between the suction pressure and the exhaust pressure is increased in proportion to the exhaust pressure so that the frictional force acting on the piston 12 is increased while the rotation speed of the piston 12 is reduced. The rotation speed of the piston 12 becomes 212 rpm in the case of 0.61 Mpa exhaust pressure and becomes 32 rpm in the case of 2.06 Mpa exhaust pressure.
When the rotation speed of the rolling piston 12 in the cylinder 13 is increased as described above, the relative sliding speed between the vane tip and the piston 12 is remarkably reduced so that the vane tip is scarcely abraded. The operational efficiency of the compressor is thus improved and life of the compressor is thus prolonged.
During exhaust of the compressed gas through the exhaust port 13b, the oil A in the shell 7 is forcibly pumped up along the oil conduit 22 of the crankshaft 1 due to centrifugal force of the rotating crankshaft 1. While the oil A is pumped up in the conduit 22, the oil A flows out through the oil ports 23, 24 and 25 of the shaft 1 and through the radial hole 11a of the eccentric sheave 11. Therefore, the oil A lubricates the contact part between the inner surface of the piston 12 and the outer surface of the sheave 11 and reduces the frictional force generated between the piston 12 and the sheave 11.
The lubrication to the contact parts between the shaft 1 and the bearing 4 and 5 is achieved by the oil supplied to those contact parts through the lubricating grooves 30 formed on the inner surfaces of the bearings 4 and 5 as shown in FIGS. 7, 8A and 8B.
However, the above compressor has a problem that the oil can not be smoothly supplied to the portion of the sheave 11 opposed to the hole 11a.
Due to the deficient oil supply for the portion opposed to the hole 11a, the frictional force between the inner surface of the rolling piston 12 and the outer surface of the sheave 11 is increased and this reduces the rotation speed of the piston 12. The frictional force between the outer surface of the piston 12 and the vane 14 is thus increased so that the vane tip is more abraded so as to reduce the operational efficiency and to shorten life of the compressor.
In addition, the lubricating grooves 30 are partially formed on the inner surfaces of the main and sub bearings 4 and 5 as shown in FIGS. 8A and 8B, it is impossible to sufficiently supply the oil for the contact parts between the shaft 1 and the bearings 4 and 5. In this regard, the outer surface of the shaft 1 will be seriously scratched.
In an effort to solve the above problems, diameters and lengths of the main and sub bearings may be reduced so as to reduce the contact area between the crankshaft and the bearings and to prevent the frictional contact between the crankshaft and the bearings. However, there is a limit in reduction of the diameters and lengths of the bearings. Even when the diameters and lengths of the bearings are fortunately reduced, the rigidity of the crankshaft will be deteriorated so that the crankshaft may be easily broken during operation of the compressor. The reduction of the diameters and lengths of the bearings inevitably results in reduction of the lubricating groove size of the inner surfaces of the bearings so that the lubricating grooves of the bearings fail to supply sufficient amount of oil and cause frictional scratches on the crankshaft.
In addition, the refrigerant may be substituted with another refrigerant in an effort to solve the above problems. However, this method is accompanied with a problem that the pressure difference between the suction chamber and the compression chamber is increased. The torque acting on the crankshaft is thus increased so that the radius of the crankshaft needs to be increased in order for keeping operational reliability of the crankshaft. However, increase of the radius of the crankshaft can not help being accompanied with mechanical loss of the shaft.